1. Field of the Invention
The present invention pertains to the field of heating, ventilating, and air conditioning ("HVAC"). More particularly, this invention relates to systems and methods for controlling the temperature and humidity of a controlled space.
2. Background
The quality of indoor air has been linked to many illnesses and has been shown to have a direct impact on worker productivity. New research strongly suggests that indoor humidity levels have a far greater impact on the health of building occupants than previously suspected. For example, microbial activity (e.g., mold and fungus), which increases at higher indoor humidity levels, has been shown to emit harmful organic compounds. Childhood asthma is now suspected by some researchers to be linked to such microbial activity.
In addition to direct health effects, the odors associated with microbial activity are often cited as a primary reason why indoor air quality is considered unacceptable to occupants. When odors are encountered in a building, building operators often respond by increasing outdoor air quantities in an attempt to eliminate odors. This often exacerbates the problem because increasing outdoor air quantities often results in higher indoor air humidity levels, which, in turn, fosters further microbial activity.
The HVAC industry has responded to these indoor air quality ("IAQ") concerns through its trade organization, the American Society of Heating and Refrigeration and Air Conditioning Engineers ("ASHRAE"). Several years ago, the ASHRAE issued IAQ Standard 62-1989, entitled, "Ventilation for Acceptable Indoor Air Quality." This standard emphasizes the need for continuous outdoor air ventilation as well as the importance of maintaining indoor humidity levels. Recently proposed modifications to the standard (62-1989R) have placed further emphasis on humidity control, specifically in hot and humid climates. The proposed standard emphasizes the need for maintaining proper humidity levels during peak and partial loading conditions, and during both occupied conditions (where it is recommended that relative humidity be maintained at no more than 60%) and unoccupied conditions (where it is recommended that relative humidity be maintained at no more than 70%). There is, therefore, a significant need for energy-efficient systems that effectively control the humidity of an indoor space while simultaneously providing high quantities of outdoor air to the space.
Facilities with high occupancy rates or high levels of ingress and egress, such as schools, hospitals, nursing homes and many offices, typically have large amounts of outdoor air introduced to the occupied space and, consequently present a significant HVAC design challenge. Due to the extreme humidity levels and large number of partial-load cooling hours that exist in hot and humid climates, maintaining relative humidity at levels recommended by ASHRAE is extremely difficult and costly if conventional HVAC approaches are to be used for such facilities.
For example, during days when the temperature is moderate but the humidity is high (partial load conditions) a packaged HVAC unit will quickly bring the space to the desired temperature, then turn off its cooling coil. As the outdoor air continues to be provided to the space, the indoor humidity level climbs until the temperature in the space once again causes the thermostat to initiate cooling operation. By this time, the mixed air condition supplied to the cooling coil is elevated in humidity. The elevated humidity level of air reaching the cooling coil results in a high dew point temperature leaving the cooling coil. The space temperature is maintained but humidity control is lost, resulting in elevated space humidity conditions.
One of the most effective design solutions for controlling the temperature and humidity of a controlled space while providing high quantities of outdoor air involves decoupling the latent load from the conventional HVAC units serving the facility. This approach allows the conventional units to be downsized to handle the sensible load only. The dehumidified outdoor air is provided directly to the controlled space via a separate system at a room neutral temperature and humidity (typically 68.degree. F. and 55 grains). This approach allows the desired quantity of outdoor air to be provided continuously while simultaneously maintaining a desirable and consistent relative humidity for the space.
Desiccant-based systems have been used to provide a continuous supply of dehumidified outdoor air. These systems can remove a significant amount of the humidity contained within the outdoor air prior to its introduction to the conventional HVAC system or, as mentioned previously, directly to the conditioned space. This allows the packaged equipment to operate as designed and to better control the space humidity despite increased outdoor air requirements.
Desiccants can be solid or liquid substances that have the ability to attract and hold relatively large quantities of water. In many commercial air conditioning applications where desiccants are used, the desiccant is in a solid form and absorbs moisture from the air to be conditioned. Examples of these types of desiccants are silica gel, activated alumina, molecular sieves, and deliquescent hygroscopic salts. In some cases, these desiccants are contained in beds over which the air to be conditioned is passed. Many times, however, the desiccant is contained in what is known as a "desiccant wheel."
A desiccant wheel is an apparatus typically comprising a plurality of closely spaced, very thin sheets of paper or metal which are coated or impregnated with a desiccant material. The wheel is frequently contained in duct work or in an air handling system that is divided into two sections. The wheel is rotated slowly on its axis such that a given zone of the wheel is sequentially exposed to the two sections. In the first section, the desiccant is contacted by the supply/outdoor air. In this section, the desiccant wheel dehumidifies the supply/outdoor air stream by absorbing moisture from the air onto its desiccant surface.
In the second section of the desiccant wheel, the desiccant contacts the return/exhaust air being discharged from the space. This return/exhaust air desorbs the moisture from the desiccant that was adsorbed from the supply/outdoor air. By cycling the wheel through these two air streams, the adsorbing/desorbing operation of the wheel is continuous and occurs simultaneously.
The prior art generally includes two types of desiccant-wheel systems: i) the "passive" dual wheel energy recovery preconditioner (DWERP) system; and ii) the "active" thermally regenerated desiccant-based cooling (DBC) system.
As shown in prior art FIGS. 1A and 1B, the DWERP system typically combines a desiccant-based total energy recovery wheel and sensible only recovery wheel along with a conventional chilled water or direct expansion ("DX") cooling coil to cool and dehumidify or heat and humidify the outdoor air supplied to a facility (depending on the ambient conditions).
A typical DWERP system operating in the cooling mode is illustrated in FIG. 1A. In the cooling mode, the supply/outdoor air stream passes through the desiccant-based total energy recovery wheel where it is precooled and dehumidified, giving up much of its humidity and heat to the desiccant coated wheel media. Next the outdoor air is passed through the cooling coil where it is further cooled and dehumidified until reaching the absolute humidity level required by the occupied space. The cold, dehumidified outdoor air stream is finally passed through a sensible only wheel where it is reheated to the desired temperature, using only the energy contained within the return air stream being exhausted from the space and passed through the secondary side of the sensible only wheel.
The return air stream leaves the sensible only wheel cool and dry, having passed through the transfer matrix cooled by the cold air leaving the cooling coil. This cool, dry air is then passed through the total energy recovery wheel where it picks up the heat and moisture contained within the desiccant coated total energy recovery matrix and exhausts it to the outdoors.
The cooling coil in the DWERP can typically be operated as low as 51.degree. F. to deliver air at an absolute humidity level of 55 grains. With custom modifications, DX cooling coils can be designed in a DWERP system to cool the outdoor air as low as 48.degree. F. to produce a delivered absolute humidity level as low as 50 grains, the practical limit for conventional cooling equipment. At an outdoor humidity condition of 130 grains (a typical dew point cooling design condition for the southern United States) the DWERP is usually capable of providing no more than an 80 grain reduction in the outdoor humidity content, resulting in a latent load reduction of approximately 544,000 BTU/hr for a 10,000 cfm system.
Referring now to prior art FIG. 2, the "active" desiccant-based cooling system (DBC) has also been applied to reduce latent loads. The "active" DBC system typically combines a thermally regenerated desiccant dehumidification wheel (an "active DH wheel"), a sensible only recovery wheel, a regeneration heat source and, in most cases, an evaporative cooler. FIG. 2 illustrates an example of a typical DBC system operating in a cooling mode. The DBC system operates by passing the outdoor/supply air through the thermally regenerated desiccant wheel, wherein the air is dehumidified and warmed. This warming occurs from both the heat of adsorption (energy released when the moisture is adsorbed on the desiccant surface) and the heat that is transferred by the dehumidification wheel matrix as it rotates from the hot regeneration air stream to the outdoor/supply air stream. Upon exiting the desiccant wheel, the hot outdoor/supply air is cooled by passing it through the sensible only wheel. At this point in a DBC system the outdoor/supply air condition is dehumidified, but at approximately the same temperature as the outdoor air. As a result, a significant amount of post cooling energy is required to reduce the supply air temperature to the desired space neutral condition.
The regeneration air stream of an active DBC system can be either a return air stream or another outdoor air stream. In either case, this air stream is usually first passed through an evaporative cooler to provide an improved driving force for the outdoor/supply air side sensible only heat recovery (in order to reduce the post cooling required). The air leaves the evaporative cooler cooled and humidified, then passes through the secondary side of the sensible only heat exchanger where the air stream is preheated prior to being introduced to the regeneration heater. Next, some or all of this air stream is passed through the regeneration heating source where it is typically heated to a temperature in the range of 175.degree. F. to 300.degree. F. This hot regeneration air stream desorbs the moisture contained in the active DH wheel's desiccant coating and exhausts it from the DBC system. The "active" DH desiccant wheel used in this DBC system must be continuously regenerated with a sizable thermal energy source in order for it to continuously dry the outdoor/supply air stream to the humidity level required by a typical application.
Even more than the DWERP approach, the active DBC system is limited in its ability to provide extremely dry air at typical dew point design conditions. The maximum moisture reduction currently available from a conventional DBC system is typically approximately 70 grains, and this level of humidity reduction generally requires the system to be operated at very low face velocities (i.e,. a very large system for the airflow processed). As a result, at an outdoor air dew point design condition of 130 grains, the delivered air condition from a DBC system will be limited to approximately 60 grains (10 grains higher than provided by the DWERP) and the temperature of the outdoor/supply air delivered to the space would usually be in excess of 80.degree.-90.degree. F.
A shortcoming of both the "passive" DWERP and the "active" DBC systems is that they cannot typically provide outdoor air drier than 50 grains and 60 grains respectively at the typical dew point design condition of 130 grains of moisture per pound of dry air. In many applications it is very desirable to provide air in the range of 35 to 45 grains so that the entire latent load can be decoupled from the conventional cooling system, allowing it to be downsized to handle only the sensible load of the building. Two application examples include schools and offices.
School facilities are often designed to trade off indoor air quality for reduced cost. For example, many school facilities recently constructed in the southern United States have been designed to provide an outdoor air volume of only 7.5 cfm/student, as opposed to the 15 cfm/student required by ASHRAE standards. Cutting the outdoor air volume in half reduces the initial cost of the facility because the size of the outdoor air preconditioning system is reduced accordingly, as is the size of the ductwork required to accommodate the outdoor and exhaust air streams. However, because the outdoor air volume is cut in half, the grain differential between that supplied to the space and that desired within the space must be doubled if all of the space latent load is to be handled by the outdoor air preconditioning system. To do so requires very dry air.
For example, a typical school classroom contains approximately 30 children. The sensible load associated with the lights, occupants and other things in the classroom is approximately 2 tons (24,000 BTU/hr.). The latent load associated with the occupants and infiltration is approximately 4.3 pounds per hour. Assuming the space is to be controlled at 75.degree. F. and 50% relative humidity, an absolute humidity content of 65 grains is desired. If the latent load is to be handled with an outdoor air load of 450 cfm (based on 15 cfm/student), then the outdoor air must be delivered at 50 grains. This can be calculated by dividing the pounds of latent load by the pounds of dry air (4.3 lbs. moisture/hr divided by 2,025 lbs. outdoor air/hour) to determine the required moisture differential required (in this case a differential of 0.0021 lb. moisture per lb. of dry air is required or 15 grains). By taking the desired space humidity content (65 grains) and subtracting the calculated moisture differential (15 grains) the required supply condition may be calculated (50 grains).
If only 7.5 cfm/student is applied, in lieu of 15 cfm, then the required moisture differential doubles from 15 grains to 30 grains. As a result, to handle the latent load with 225 cfm of outdoor (7.5 cfm/student) the outdoor air must now be delivered at 35 grains (65 grains -30 grain moisture differential). As mentioned previously, this is much drier than can be provided with conventional DWERP or DBC systems.
A similar scenario exists in a typical office environment where the required 20 cfm per person comprises only approximately 20% outdoor air. For example, the office building may require 20,000 cfm of supply air of which 4,000 is outdoor air. If the outdoor air could be dehumidified to 40 grains by an outdoor air preconditioning system, then this air, when mixed with 16,000 cfm of return air at 65 grains would require the air entering the space to have an absolute humidity level of 60 grains to maintain the desired space relative humidity of 50% at 75.degree. F. A significant benefit is achieved because the cooling coil can now be operated to provide air at only 58.degree. F. to handle the sensible load as opposed to 52.degree. F. that would have been required to deliver the 60 grain air if it had not been dehumidified by the outdoor air preconditioner. This increased dehumidification capability would have a very positive impact on the cost of the project, chiller efficiency and energy consumption, and would improved humidity control during unoccupied conditions.
Conventional DWERP and DBC dual wheel systems both lack the capability to efficiently provide very dry air (35 to 45 grains when processing outdoor air at typical dew point design conditions 120 to 130 grains) that is often necessary to control the space relative humidity below 60% as recommended by ASHRAE. Hence, there is a strong need for an improved outdoor air preconditioning system for controlling both the temperature and humidity of occupied and unoccupied spaces that is capable of delivering air at very low absolute humidity levels (below approximately 48 grains), in an energy efficient manner.
Another disadvantage of DBC active desiccant dual wheel systems is that they require an external heat source to regenerate the desiccant wheel to drive the dehumidification process. Such systems also typically apply an evaporative cooling section to help remove much of the heat generated by the desiccant wheel as part of the adsorption process. This heat of adsorption and carry-over heat from the desiccant wheel significantly raises the temperature of the outdoor air stream. The portion of the heat that is not removed by the second, sensible only wheel and evaporative cooling section must be removed using post-cooling energy from either in the DBC system or added to the HVAC systems serving the space. The requirement for regeneration heat adds complexity to the system's operation, installation and control sequence, and often requires a different heating source than that used for space heating because of the high temperature required to dehumidify humid outdoor air to the desired humidity level. The evaporative cooling section increases system maintenance, such as water treatment and winterization, and often increases microbial growth. Thus, there is also a strong need for an outdoor air preconditioning system that can deliver very dry air without the use of a regeneration heating source or an evaporative cooling section.
Another shortcoming of conventional DBC active desiccant dual wheel systems is limited flexibility. These systems can provide warm dry air. But, as the outdoor air humidity increases, the outlet temperature from the unit also increases. As a result, the DBC system cannot provide cool dry air (without significant post-process cooling) when the space conditions are hot and humid. Thus, it would be advantageous to provide a system that could provide dry, cool air when the space becomes hot and humid, thereby having the flexibility to handle various combinations of indoor and outdoor temperature and humidity conditions.
For the foregoing reasons there is a strong need for an energy efficient system for controlling the temperature and humidity level of the air of a controlled space that is capable of delivering preconditioned outdoor air at very low absolute humidity levels without requiring an external heat source and that has the flexibility to handle various combinations of temperature and humidity. The present invention provides these and other advantageous results.